Earth boring bit with two piece bearing and rigid face seal assembly

ABSTRACT

An earth boring bit having a cantilevered bearing shaft and a compensator system to equalize the pressure of the lubricant with the hydrostatic pressure of the drilling fluid surrounding the bit. A rigid face seal assembly, positioned between the cutter and bearing shaft of the bit, moves axially in response to, and to compensate for, dynamic pressure changes in the lubricant adjacent the seal. The positioning and sizing of resilient energizer rings in relationship to the geometries of the mating grooves between the cutter and shaft and the rigid sealing rings of the face seal assembly minimize axial seal movement relative to cutter and shaft during drilling to enhance seal life. A threaded bearing lug receives an internally threaded bearing sleeve. The journal bearing surface is approximately midway between the shaft seal groove, a portion of which is on its bearing sleeve and another portion on the cutter seal groove.

CROSS REFERENCE TO RELATED PATENTS

This application is related to, and is a continuation-in-part of, aprevious application: "Earth Boring Bit With Improved Rigid Face SealAssembly" Ser. No. 733,435, filed May 13, 1985, now issued as U.S. Pat.No. 4,666,001, which is a continuation-in-part of application, "EarthBoring Bit With Pressure Compensating Rigid Face Seal", Ser. No. 542,801filed Oct. 17, 1983, now U.S. Pat. No. 4,516,641.

This application has disclosure in common with an application of JosephL. Kelly, Jr., entitled "VOLUME AND PRESSURE BALANCED RIGID FACE SEALFOR ROCK BITS", Ser. No. 023,178, filed Mar. 9, 1987.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates to earth boring bits, lubricated with a systemwhich includes a hydrostatic pressure compensator to balance theinternal pressure of the lubricant inside the bit with the hydrostaticpressure of a liquid drilling fluid that surrounds the bit duringdrilling. In this combination, the specific improvement relates to theseal and bearing assembly between each cutter and bearing shaft.

2. Background Information

The preferred embodiment in the above U.S. Pat. No. 4,516,641 utilizes apair of rigid face seals positioned in a seal groove, including a pairof resilient energizer rings, preferably of the O-ring type. Thedimensional relationships of the sealing components and the grooveresult in greater axial movement of the rigid face seals than theassociated cutter.

As a bit rotated during drilling, its cutters move axially, or with arocking motion, on the bearing shafts because of the clearances andnormal manufacturing tolerances. Some clearances are necessary toassemble the cutters on the shafts. Axial and radial cutter movementswhich results from the clearance between cutter and shaft causes rapidpressure variations in the lubricant, or more accurately, volume changesin the lubricant in the vicinity of each seal. In the preferredembodiment of the above U.S. Pat. No. 4,516,641, the rigid face sealsmay move axially a distance greater than the axial cutter movement by aratio of about 1.88 to one.

The disclosure in the second of the above U.S. Pat. No. 4,666,001teaches the positioning of the seal groove and seal assembly in relationto the bearing surface such that rigid face seal movement is decreasedrelative to axial cutter movement during drilling to enhance seal life.Preferably the ratio of rigid ring movement to axial cutter movement issubstantially one-half to one.

A two piece bearing construction is disclosed in U.S. Pat. No.4,600,064, "Earth Boring Bit With Bearing Sleeve", July 15, 1986. Aninternally tapered and threaded bearing sleeve is made up on a mating,externally threaded bearing lug. The mouth of the sleeve engages ashoulder on the base region of the the bearing lug and has a selectedradial thickness such that the sleeve may be made up to a selectedtorque. Also, the bearing sleeve has a length greater than that of thethreaded portion to define a thick walled inner end region to receive aresilient retainer ring in a groove that provides a selected sectionover the threads on the interior of the sleeve.

SUMMARY OF THE INVENTION

The general object of this invention is to provide an improved rigidface seal assembly in a rock bit of the type susceptible to pressure orvolume changes in the vicinity of the seal assembly when the associatedcutter moves during drilling. The improvement decreases the pressure orvolume changes in the vicinity of the seal assembly by positioning theseal groove and seal assembly in relation to the bearing surface suchthat rigid face seal movement is decreased relative to axial cuttermovement. In a preferred embodiment the axial movement of the rigid faceseal with respect to the shaft is about one half that of the cutter, andis achieved by making the bearing surface intermediate the radialthickness of the seal assembly groove. That is, part of the groove is inthe cutter and the other is in the bearing shaft. The relationship ofseal movement to cutter movement is established by dimensioningconsistent with the following formula, derived for this invention:

    S=2G-C/H+C+2A

Where:

S=Seal Movement For Unit Cone Movement Relative to the Shaft

H=Effective Shaft O-Ring Annular Area

C=Effective Cutter O-Ring Annular Area

A=Effective Rigid Ring Annular Area

G=Effective Cutter Seal Groove Annular Area

The preferred embodiment uses a threaded bearing lug upon which isthreaded a bearing sleeve containing about one half the radial thicknessof the total seal groove. This portion of the seal groove is called theshaft groove. The other radial portion of the total seal groove isformed in the cutter.

Additional objects, features and advantages of the invention will becomeapparent in the following description.

DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view in longitudinal section of a portion of an earth boringbit, showing the compensator system, bearing shaft, cutter and oneembodiment of a seal assembly.

FIG. 2 is a fragmentary view in longitudinal section of the lowerportion of a bit, enlarged with respect to FIG. 1, to better expose theseal assembly.

FIG. 3 is a fragmentary view in longitudinal section of yet anotherportion of the cutter and bearing shaft, showing the bearing sealassembly enlarged with respect to the illustration of FIG. 2.

FIG. 4 is an enlarged longitudinal section of one of the rigid rings ofthe seal assembly.

FIG. 5 is a fragmentary, longitudinal section of the seal seat andconical surface seal groove in the cutter.

FIG. 6 is a longitudinal section of the seal seat annular insert used onthe bearing shaft to form the conical, contoured surface that receivesand positions the seal assembly.

FIG. 7 is a fragmentary view in longitudinal section of the lowerportion of an alternate embodiment of the invention shown in FIG. 1.

FIG. 8 is a view in longitudinal section of a portion of an earth boringbit having an improved rigid face seal assembly to reduce duringdrilling the amount of seal movement relative to cutter movement.

FIG. 9 is an enlarged fragmentary view in longitudinal section of thepreferred seal assembly and groove embodiment shown in FIG. 8.

FIG. 10 is a fragmentary view of another embodiment of an earth boringbit, partially in longitudinal section, that utilizes a threaded bearingsleeve to form the journal bearing of the shaft and a portion of a sealgroove.

FIG. 11 is an enlarged fragmentary view in longitudinal section of thebearing, cutter and seal assembly of FIG. 10.

DESCRIPTION OF THE PREFERRED EMBODIMENTS FIGS. 1-7

The improvement of this invention is best appreciated by initialreference to the disclosure of one of the related patents, U.S. Pat. No.4,516,641, which is shown in FIGS. 1-7 herein. The numeral 11 in FIG. 1designates a lubricated, rotatable cone or cutter type earth boring bithaving a body formed in three head sections or legs 13, only one ofwhich is shown. Each leg 13 includes an oblique cantilevered bearingshaft 15 that depends inwardly and downwardly to support a rotatablecutter 17 having earth disintegrating teeth 19. Lubricant passage 21supplies lubricant to the bearing surfaces between the bearing shaft 15and cutter 17. A seal assembly 23 retains lubricant in the bearing andprevents borehole fluid from entering the bearing. A hydrostaticpressure compensator is part of a lubrication system 25 connected withthe lubricant passage 21 to equalize the pressure of the liquidlubricant inside the bearing with the hydrostatic pressure of the fluidin the borehole. A preferred compensator system is shown in the patentto Stuart C. Millsapps, Jr., U.S. Pat. No. 4,276,946.

The geometry of the bearings on the shaft 15 and within the cutter 17are of a prior art configuration, including the use of a ball bearingretainer 27, which with a plug 26 welded at 28 retains the cutter on thebearing shaft, preferably as shown in the U.S. patent of Robert A.Cunningham, U.S. Pat. No. Re. 28,625.

Referring especially to FIG. 3, the cutter and shaft include an annularseal groove or gland that has axially spaced, generally radial end walls29, 31 and inner and outer circumferential walls 33, 35. End wall 31 andcircumferential wall 33 are formed upon a seal seat insert 36 secured tothe bearing shaft 15.

The seal assembly includes a pair of annular rigid rings 37, 39 withopposed radial faces 41, 43. The pair of rigid, preferably metal, ringshave a radially measured thickness less than the annular space betweenthe inner and outer side walls 33, 35 of the groove and an axiallymeasured width which is less than the width or the distance between theend walls 29, 31 of the groove, as will be explained more fullyhereafter.

Each of a pair of resilient, energizer rings 45 or 47 extends between aseal seat 49 or 51 on one of the metal rings 37 or 39 and an opposedseal seat 53 or 55 on the inner or outer circumferential walls 33, 35.Each seal seat has an annular groove and configuration to position andretain the associated energizer ring and the metal ring, which issuspended within the groove intermediate the circumferential walls 33,35 and the end walls 29, 31 to provide clearances C₁ and C₂ which existwhen the thrust surfaces 32 are in contact. The positions of seats 53and 55 relative to each other are selected such that, at assembly, theinitial deflection of each seal half relative to its adjacent end wall29 or 31 will provide sufficient contact pressure between radial faces41 and 43 to maintain sealing contact between all the elements of theseal assembly throughout the full range of seal movements permitted byclearances C₁ and C₂ and the play between cutter and shaft. See U.S.Pat. No. 3,180,648 for a description of an earlier construction of aconical, "Duo Cone" seal arrangement of the Caterpillar Tractor Company,and U.S. Pat. Nos. 3,403,916, 3,524,654 and 4,077,634 for improvementsto such seals.

From FIG. 3 it may be seen that one of the metal rings 37, 39 isinverted with respect to the other, a feature which permits the sealassembly to span the groove diagonally and engage oppositecircumferential walls 33, 35. The clearances C₁ and C₂ are between eachof the end walls 29, 31 of the groove and the engaged rigid rings 37,39. Drilling fluid fills the space 57 and acts upon the outermost sideof the seal assembly 23, and lubricant fills the space 29 and acts uponthe innermost side of the assembly. The rigid rings 37 or 39 have abeveled, substantially conical portion on the lubricant side of the sealassembly to define a space 61 to feed lubricant to the engaged radialfaces 41, 43, which regenerate inwardly as they wear in service. SeeU.S. Pat. No. 3,180,648 for a description of one configuration of suchseal faces.

The dimensions provided below relate to the bit used in the first fieldtest of the invention, which was a "Hughes" 121/4 inch, J22 type bit.With reference especially to FIGS. 4 and 5, the radial thickness T ofeach of the three metal rings 37 was 0.200 inch, the axial width W wasabout 0.270 inch and the outside diameter was about 3.449 inches. Angleα was about twenty degrees and the radii R₁ and R₂ averaged 0.048 and0.080 inch respectively, R₁ being tangent with the conical surface 63.The axial dimensions Y and Z averaged respectively 0.050 and 0.149 inch.The depth D of the positioner groove 65 below the lip 67 averaged 0.009inch to help position and confine the associated energizer ring 45during assembly. The radial thickness X of the radial sealing face 41was about 0.050 inch with a surface finish of about one or two RMS and atapered surface defined by a spherical radius R₃ of about 80 inches witha surface finish of about two or three RMS.

The inverted and opposing metal rings 39 has a radial thickness T ofabout 0.199 inch, a radial width W of about 0.247 inch and an outsidediameter of about 3.450 inches. Angle α was about 19 degrees, and theradii R₁ and R₂ were both about 0.075 inch. The axial dimensions Y and Zwere respectively about 0.023 and 0.154 inch and the depth D of thepositioner groove being about 0.016 inch. There was a flat sealing faceon the rings 39 that extended across the entire thickness T of the ring,that had a surface finish similar to that of rings 37.

Ring 37 was purchased from Caterpillar Tractor Company and is one oftheir standard hard metal alloy rings. Ring 39 was manufactured byHughes Tool Company specifically for this invention from an airhardening tool steel.

The configuration of the seal seat in the cutter 17 of the bit is shownin FIG. 5. The seal seat was defined by a conical surface 70 having anangle θ₁ of about 191/2 degrees, a positioner groove or seal seat 53having a radius R₄ of about 0.060 inch located a distance D₂ of about0.129 inch from end wall 29, and a depth D₁ of about 0.008 inch. Theconical surface 70 intersected the groove 53 at a point 72 which waslocated radially from the surface 35 a distance of about 0.021 inch.

A similar configuration was used for the seal seat on the bearing shaft15, defined in this instance by the seal seat insert 36 shown in FIG. 6having a thickness T₁ of 0.105 inch. The positioner groove had a depthD₃ of about 0.011 inch, formed by a radius R₅ of about 0.060 inchlocated a distance D₄ about 0.140 inch from end wall 31. θ₂ was aconical angle of about 20 degrees located in a manner similar to theconical surface 70 of FIG. 5.

The O-rings or energizer rings 45, 47 after service had across-sectional thickness of about 0.168 inch, a hardness of about 59durometer, Shore A, inside diameters of about 3.057 and 2.760 inchesrespectively, and a high resilience, measured to be about 43 percentrebound using the above described O-rings and a Shore Resiliometer,Model SR-1. The radial end walls 29, 31 of the seal groove were locateda width of about 0.580 inch apart with bearing thrust surfaces 32 inmutual contact. Using the above components, the assembly loading on thefaces of the rigid rings was about 40 to 60 pounds, as determined fromload deflection curves. The clearances C₁ and C₂ were respectively about0.035 and 0.029 inch at assembly with thrust surfaces 32 in contact todefine the minimum groove width and the diameters of the circumferentialwalls 33, 35 were respectively about 2.969 and 3.529 inches.

For the first bit tested, the axial bearing play of each of the cutters,after testing was:

Axial Play (inch)

No. 1 Cutter: 0.012

No. 2 Cutter: 0.015

No. 3 Cutter: 0.012

In operation, and during drilling in a well bore filled with liquid, thecompensator 25 acts to balance the hydrostatic pressure of the liquid inthe well bore with the pressure of the liquid lubricant inside thebearing. However, cutter movements during drilling, caused by thecomplex forces applied to a cutter, and the clearances which are ofnecessity used to enable assembly of the parts, produce rapid changes inthe volume defined by the space 59. The viscosity of the lubricant andflow restrictions between the space 59 and hydrostatic pressurecompensator 59 do not allow compensation for the volume changes in space59 as rapidly as they occur. Nevertheless, seal assembly 23 will movesufficiently to provide the required volume change and thereby minimizethe pressure changes experienced by the seal which would otherwise causerapid depletion of the lubricant supply or entry of borehole fluids intothe bearing, with resulting bearing and seal damage.

Use of the seal assembly 23 described above in a bit which includes ahydrostatic pressure compensator minimizes the pressure differentials towhich the seal assembly is exposed through the cooperative relationshipof the hydrostatic pressure compensator and the dynamic pressurecompensating abilities of the seal assembly. The seal assembly is onewhich spans diagonally the seal groove such that one of the resilientenergizer rings engages a wall of the cutter, while the other energizerring engages a wall of the shaft. Thus, the outermost portion of each ofthe energizer rings is exposed to the fluid in the borehole, while aninnermost portion of each of the energizer rings is exposed to thelubricant inside the bearing. Every pressure differential is thereforesensed by the seal assembly, which is moved by each such pressuredifferential. A seal assembly which cannot be moved by the differentialpressure cannot effectively compensate for dynamic changes in the volumeof space 59. Preferably the seal half consisting of energizer ring 47and rigid ring 39 should have the same axial load deflectioncharacteristics as the mating half consisting of energizer ring 45 andrigid ring 37 to balance and minimize the increase in the loading ofengaged radial faces 41 and 43 caused by pressure differentials.

Another requirement for a satisfactory seal assembly, and the groove inwhich it is placed, is that the assembly be positioned between the endwalls of the groove to permit unrestricted axial movement of the rigidrings between the walls of the groove in response to sensed pressuredifferentials. If the bearing lubricant could freely enter and leavespace 59 as the volume of space 59 is changed by cutter movement, thepressure differentials acting on the seal would be negligible and themovement of the rigid rings would be less than the cutter movement.Furthermore, if the load deflection characteristics of each half wereequal, as preferred, the rigid ring movement would be on-half the cuttermovement in the above described embodiment. However, because lubricantmovement is restricted, greater rigid ring movement must be providedfor. The required clearances C₁ and C₂ were determined by building amodel of the seal cutter and shaft assembly and measuring the movementof the rigid rings in response to cutter movement with the exit fromspace 59 blocked; for example, by a conventional O-ring seal between thebearing surfaces of the simulated cutter and shaft. To be sure thataccurate rigid ring movement takes place in the model, it is importantto have space 59 completely filled with an incompressible fluid that isfree of any air or vapor pockets. Furthermore, in some cases, it may benecessary to pressurize space 57 with air to insure complete rigid ringmovement in response to movement of end wall 29 away from end wall 31.

A model as hereinabove described was used to measure the movement of therigid rings in response to cutter movement for the shaft, cutter andseal used in the first test bit. Air pressure in space 57 was notrequired for this test because the pressure in space 59 did not dropbelow the vapor pressure of the fluid used to fill the space. The ratioof the rigid ring movement to cutter movement was determined from themeasurements to be 1.88:1. This ratio is influenced by the geometry ofspace 59, the size, shape and elastic properties of the energizer ringsand the manner in which the energizer rings are deformed by the rigidrings and wall of the seal groove. Thus, a change in any of theseparameters is likely to cause a change in required clearances C₁ and C₂.

After the ratio of the rigid ring movement to cutter movement has beenestablished, as described above, the minimum values for C₁ and C₂ may becalculated. The maximum seal or rigid ring movement with respect to thebearing shaft is calculated by multiplying the axial play between thecutter and shaft by the ratio of rigid ring movement to cutter movement.When bearing thrust surfaces 32 are in contact, the first axialclearance C₁ between rigid ring 37 and the inner wall 29 of the grooveshould be greater than the maximum rigid ring movement less the axialplay between the cutter and the shaft. The second axial clearance C₂between rigid ring 39 and the outer end wall 31 of the groove measuredwith the thrust surfaces 32 in contact should be greater than a valueequal to the maximum rigid ring movement as calculated above less thedisplacement which rigid rings 37 and 39 undergo when the space betweenend walls 29 and 31 is increased by axial play from its minimum lengthto its maximum length in the absence of any pressure differential acrossthe seal. This displacement of rigid rings 37 and 39 in the absence of apressure differential can be determined with the model if space 59 isvented or it can be calculated with the aid of the load deflectioncurves for the seal halves.

While the embodiment of the invention disclosed above was that of theinitial test, the commercial embodiment is expected to be closer to thatshown in FIG. 7. The leg 101 includes an oblique cantilevered bearingshaft 103 that depends inwardly and downwardly to support a rotatablecutter 105 having earth disintegrating teeth 107. A lubricant passage109 supplies lubricant to the bearing surfaces between the bearing shaft103 and the cutter 105. A snap ring retainer 106 similar to that shownin U.S. Pat. No. 4,344,658 is used in place of the ball retainer shownin FIG. 1.

A seal assembly 111 retains lubricant and excludes borehole fluids. Thisseal assembly has the same configuration as assembly 23 of FIG. 3,however, the innermost energizer ring 113 engages directly the journalbearing cylindrical surface 115, rather than a seal seat insert 36. Aseal seat configuration is provided similar to the seal seat 55 andinner circumferential wall 33 in the FIG. 3 embodiment. This reduces thediameter of the seal seat of FIG. 7, as compared to the diameter of theseal seat in FIG. 3. This reduction in diameter of the seal seat inrelation to the diameter of journal bearing cylindrical surface 115reduces the ratio of rigid ring movement to cutter movement. This ratio,determined by making a model similar to the one described in connectionwith the FIG. 3 embodiment, except using the FIG. 7 bearingconfiguration, is 1.28:1. The materials for the various components ofthe seal assembly are identical with the materials used in theembodiment of FIGS. 1-4 except both rigid rings are preferablyconstructed of the same hard metal alloy at ring 37.

FIGS. 8-9

An alternate to the disclosure above appears in the application fromwhich this application continues, now U.S. Pat. No. 4,660,001, and isshown in FIGS. 8 and 9.

In FIG. 8 the numeral 121 designates one section of a rock bit having athreaded upper end or shank 123, a lubrication system 125 of the typepreviously described, which feeds lubricant a the passage 127, and to apassage 129 formed in a ball plug 131 secured by weld 133 to the section121. Additional passage means 134 enables lubricant communication withthe bearing shaft 135, which has on its interior and cantilevered end apilot pin 137 and ball bearing raceway 139.

The ball plug 131 retains plural ball bearings 143 in the ball bearingraceway 139 and in a mating raceway 145 in the cutter 147. A bearingsleeve 149 has an interior surface 151 which engages with interferencefit a mating portion on the bearing shaft 135. An annular exteriorbearing surface 175 engages a mating cylindrical bearing surface 173 inthe cutter 147. These bearing elements enable rotation of the cutter 147such that the earth disintegrating teeth 153 engage and disintegrategeological formations during drilling.

To seal lubricant within the cutter 147 for lubrication of the abovedescribed bearing surfaces, a seal assembly 155 is provided, which maybe better seen with reference to FIG. 9.

The seal assembly 155 is disposed partially within a generally L shapedgroove in the shaft 135, as best seen in FIG. 9, having a radial endwall 157 and a circumferential wall 159. Another groove is locatedwithin the cutter 147, having a radial end wall 161 and circumferentialwall 163.

Confined within the grooves are a pair of energizer rings 165, 167 andrigid annular rings 169, 171. The geometries of the groove surfaces, theenergizer rings and the annular rings shown here are identical withthose described in the preceding embodiment. However, the bearingsurface 175 is located substantially midway between the circumferentialwalls 159, 163 and is defined by a bearing sleeve 149 having an annularsurface 151 retained by interference fit in a mating annular slot with asmall shoulder 177 located in the bearing shaft.

The improvement of FIGS. 8 and 9 decreases the lubricant pressure andvolume changes in the vicinity of the seal assembly by positioning theseal groove and seal assembly in relation to the bearing surface 175such that the ratio of rigid face seal movement to cutter movement isabout one half to one. This is achieved by making the bearing surface175 intermediate the radial thickness of the seal assembly grooves orthe diameters of the circumferential surfaces 159, 163. That is, part ofthe groove is in the cutter and the other is in the bearing shaft. Therelationship of seal assembly movement to cutter movement is establishedby dimensioning consistent with the following formula, derived for thisinvention:

    S=2G-C/H+C+2A

Where:

S=Seal Movement For Unit Cone Movement Relative to Shaft

H=Effective Shaft O-Ring 165 Annular Area

C=Effective Cutter O-Ring 167 Annular Area

G=Effective Cutter Seal Groove Annular Area

The ratio of seal movement to cutter movement of about one half to oneis achieved using the above formula, a circumferential wall 159 diameterof 2.970 inch, and the following annular areas:

H=0.945 sq. inch

C=1.041 sq. inch

A=0.587 sq. inch

G=1.311 sq. inch

The formula assumes no radial bearing clearance, no lubricant flow to orfrom the seal assembly, no or off-setting changes in annular areas H, C,A and G, and rolling, nonsliding movement of the energizer rings 165,167.

The preferred assembly method uses the steps of forming about one halfthe groove in the cutter 147 and the other half in the shaft 135,assembling one rigid ring 169 and energizer ring 165 in the seal regionof the shaft 135, assembling the sleeve 149 and then the other rigidring 171 and the mating energizer ring 167 in the cutter 147, and thenassembling the cutter 147 and sleeve 149 on the shaft 135 againstshoulder 177, the sleeve 149 engaging the shaft 135 with an interferencefit.

During drilling with the embodiment shown in FIGS. 8 and 9, cutter 147will move on the shaft 135 in a complex, wobbling fashion. The aboveformula assumes simpler, axial movements of the cutter but is thought tobe sufficiently accurate to be useful in establishing design parametersfor groove and seal assembly geometries.

The advantages of the embodiment of FIGS. 8 and 9 over the embodiment ofFIGS. 1 through 7 are: (1) Pressure changes in the lubricant adjacentthe seal assembly during drilling and consequent cutter movement aregreatly reduced, ideally eliminated; (2) variations in loading on thefaces of the rigid rings 169, 171 are minimized; (3) rolling motions ofthe energizer rings 165,167 are more equalized; (4) since any axialmovement of the seal assembly is less than axial movement of the cutter,the clearances with end walls 157, 161 may be minimized.

FIGS. 10-11

The improvement of this continuation-in-part is disclosed in FIGS.10-11. Here, an earth boring bit 201 has three head sections 203 thatare welded to form a body. Extending inwardly and downwardly from eachof the sections 203 is a cantilevered bearing lug 205, threaded at 207to receive an internally threaded bearing sleeve 209. This type ofbearing arrangement is shown in U.S. Pat. No. 4,600,064, "Earth BoringBit With Bearing Sleeve", July 15, 1986.

In an upper portion of each head section 203 is a compensator 211 thatforms part of a compensator system that includes passages 213, apreferred form of compensator being shown in U.S. Pat. No. 4,276,946,"Biased Lubricant Compensator For An Earth Boring Drill Bit", July 7,1981. Conventional nozzle means 214 direct drilling fluid toward aborehole bottom. A part of the passage 213 extends through the leg 215of the section 203 into intersection with another passage 217, formed inthis instance coaxially with the bearing lug 205. Lubricant isintroduced from passage 217 through a passage 218 in a pilot pin 220formed on sleeve 209.

The thread 207 formed on the exterior of the bearing lug 205 divergesoutwardly from an inner end region 219 to an outer base region 221,having a shoulder 223 which is transverse or perpendicular with respectto the rotational axis of the lug 205. The thread form is an A.P.IV-0.040 with a 2.708 inch pitch diameter at 0.625 inch from shoulder223.

The bearing sleeve 209 attached to the bearing lub 205 has asubstantially cylindrical exterior journal bearing surface 225, atapered and threaded interior portion 227 that mates with the threads207 of the lug 205, and a traverse mouth 229 that mates with theshoulder 223 of the base region 221 of the lug. The radial thickness ofthe mouth 229 is substantially 0.345 inch, for a 143/4 inch bit providedby way of example, the sleeve beign made from a metal alloy with aminimum yield strength so as to provide a torsional yield strength ofsubstantially 11,000 foot pounds.

From the drawing it is apparent that the bearing sleeve 209 and itscylindrical journal bearing portion 225 has a length greater than thatof its threaded portion 227 to define a thick walled inner end portion231 in which is formed an assembly groove 233 that opposes or registerswith a retainer groove 235 formed in the cylindrical bearing portion 237of the cutter 239. The minimum thickness of the metal between theassembly groove 233 and the threaded portion 227 of the bearing sleeve209 for the 143/4 diameter bit provided by way of example should besubstantially 0.422 inch. A drive pin hole 241 provides a means to applya selected torque of about 2500 foot pounds to the sleeve on assemblywith the lug 205. The cutter is of a conventional configuration, withearth disintegrating teeth 243, and is retained on the bearign sleeve209 with a resilient snap ring 245 having a curved cross section andgroove configuration with curved bottom portion of the type disclosed inU.S. Pat. No. 4,236,764, "Earth Boring Drill Bit With Snap Ring CutterRetention", Dec. 2, 1980.

The sleeve 209 has a boronizing treatment of the type describe in U.S.Pat. No. 3,922,038, "Wear Resistant Boronized Surfaces And BoronizingMethods", Nov. 25, 1975, on the exterior cylindrical surface 225 toimprove wear resistance. This treatment provides the requisitieimprovement to wear resistance without causing a substantial weakeningof the sleeve.

A bearing seal assembly 247 is essentially that as shown in connectionwith the above description of FIGS. 1-7 but the seal groove arrangementis similar to that of FIGS. 8-9. Referring to FIG. 11, the seal assembly247 is disposed partially within a generally L shaped, shaft seal groove249, a portion of which is in the outer end of the sleeve 209. The shaftseal groove 249 includes a circumferential, generally cylindrical outerwall 251 and a radial inner end wall 253. A second portion 255 of theshaft seal groove 249 includes a circumferential wall 257 and a radialouter end wall 259 formed on the shoulder 223 on the base region of thebearing lug 205. The mouth 229 of the sleeve 209 is transverse andsealingly engages an opposed surface 263 of the shoulder 223 of thebearing lug 205.

An opposed, generally L-shaped cutter seal groove 265 is formed in theouter, mouth region 267 of the cutter 239, and includes acircumferential, generally cylindrical wall 269 and a radial end wall271.

Confined within the above described cutter seal groove 265 and shaftseal groove 249 are a pair of energizer rings 273, 275 and a pair ofrigid annular rings 277, 279. The geometries of the groove surfaces, theenergizer rings and the annular rings shown here are identical withthose described in connection with FIGS. 8-9. The journal bearingsurface 225 of the bearing sleeve is located substantially midway betwenthe circumferential walls 251, 269 in the preferred embodiment.

The improvement of FIGS. 10-11 decreases the lubricant pressure andvolume changes in the vicinity of the seal assembly by positioning theseal groove and seal assembly in relation to the journal bearing surfacesuch that the ratio of rigid ring seal movement to cutter movement isabout one half. Seal movement for unit cone movement is established byusing dimensions A, C, G and H consistent with the formula provided indescribing FIGS. 8-9.

The advantages of the embodiment of FIGS. 10-11 are similar to those ofFIGS. 1-7 and 8-9, but in addition, the assembly problems associatedwith FIGS. 8-9 are avoided.

While the invention has been shown in only the preferred forms, itshould be apparent to those skilled in the art aht it is not thuslimited, but is susceptible to various changes and modifications withoutdeparting from the spirit thereof.

I claim:
 1. An earth boring bit with an improved pressure compensating,rigid face seal assembly having one rigid ring and sealing face carriedby a bearing shaft and an opposed sealing face carried by the cutter,said bit comprising:a body; a threaded and cantilevered bearing lugpositioned to extend obliquely inwardly and downwardly from the body; aninternally threaded bearing sleeve screwed onto the bearing lug to forma journal bearing on the bearing shaft; a cutter having a generallycylindrical bearing formed internally therein, secured for rotationabout the journal bearing of the shaft; a lubrication system in thebody, including a hydrostatic pressure compensator, to lubricate saidbearings; a cutter seal groove formed near the outermost region of thecylindrical bearing in the cutter to have a circumferential, generallycylindrical wall; a shaft seal groove formed at least partially in thebearing sleeve to oppose the cutter seal groove and having acircumferential, generally cylindrical wall; at least one rigid ringpositioned between the shaft and cutter seal grooves with one sealingface engaging the opposed sealing face carried by a selected one of theshaft and the cutter; at least one resilient energizer ring sealinglyengaging a selected circumferential wall of the cutter and shaft sealgroove and sealingly engaging the rigid ring; the diameter of thejournal bearing being intermediate the circumferential walls of thecutter and shaft seal grooves to decrease axial movement of the rigidring with respect to axial cutter movement to a selected ration duringdrilling.
 2. The invention defined by claim 1 wherein the ratio of theseal movement to axial cutter movement is selected to decrease sealmovement, consistent with the following formula:

    S=2G-C/H+C+2A

Where: S=Seal Movement For Unit Cutter movement Relative to the ShaftH=Effective Shaft O-Ring Annular Area C=Effective Cutter O-Ring AnnularArea A=Effective Rigid Ring Annular Area G=Effective Cutter Seal GrooveAnnular Area.
 3. The invention defined by claim 1 wherein the ratio ofrigid ring movement to axial cutter movement is substantially one halfto one.
 4. An earth boring bit with an improved pressure compensating,rigid face seal assembly having one rigid ring and sealing face carriedby a bearing shaft and an opposed sealing face carried by the cutter,said bit comprising:a body; a threaded and cantilevered bearing lugpositioned to extend obliquely inwardly and downwardly from the body; aninternally threaded bearing sleeve screwed onto the bearing lug to forma generally cylindrical journal bearing on the bearing shaft; a cutterhaving a generally cylindrical bearing formed internally therein,secured for rotation about the journal bearing of the shaft; alubrication system in the body, including a hydrostatic pressurecompensator, to lubricate said bearings; a cutter seal groove formednear the outermost region of the cylindrical bearing in the cutter tohave a circumferential, generally cylindrical wall; a shaft seal grooveformed at least partially in the bearing sleeve to oppose the cutterseal groove and having a circumferential, generally cylindrical wall; apair of rigid rings positioned in the seal grooves to have opposedsealing faces; a pair of resilient energizer rings, eash of whichsealingly engages a respective one of the rigid rings and one of theoppositely facing circumferential walls of the cutter and shaft sealgrooves to define a real assembly; the diameter of the journal bearingbeing intermediate the diameters of the circumferential walls of thecutter and shaft seal grooves to decrease axial movement of the rigidrings with respect to axial cutter movement to a selected ration whenthe cutter moves axially on the bearing shaft during drilling.
 5. Theinvention defined by claim 4 wherein the ratio of the seal movement toaxial cutter movement is selected to decrease seal movement, consistentwith the following formula:

    S=2G-C/H+C+2A

Where: S=Seal Movement For Unit Cutter Movement Relative to the ShaftC=Effective Cutter O-Ring Annular Area A=Effective Rigid Ring AnnularArea G=Effective Cutter Seal Groove Annular Area.
 6. The inventiondefined by claim 4 wherein the ratio of rigid ring movement to axialcutter movement is substantially one half to one.
 7. An earth boring bitwith an improved pressure compensating, rigid face seal assembly havingone rigid sealing face carried by a bearing shaft and an opposed sealingface carried by the cutter, said bit comprising:a body which includes atleast one leg and a cantilevered and threaded bearing lug that extendsdownwardly and inwardly; a lubrication system formed at least partiallyin the body, including a hydrostatic pressure compensator, to lubricatesaid bearings; the bearing lug including a threaded inner end and anouter base region with a transverse shoulder secured to the body; abearing sleeve with a substantially cylindrical journal bearing surfacecontaining an assembly groove for a cutter retainer means, a threadedinterior secured to the the threaded end of the bearing lug, and a mouthengaging the transverse shoulder of the lug; a rotatable cutter havingan interior cylindrical bearing surface, a retainer groove thatregisters with the assembly groove in the bearing sleeve, and anannular, cutter seal groove in the outer end of the bearing surface;retainer means between the assembly groove and the retainer groove; acutter seal groove formed near the outermost region of the cylindricalbearing in the cutter to have a circumferential, generally cylindricalwall and a generally radial end wall; a shaft real groove formed atleast partially in the outer end of the bearing sleeve and thetransverse shoulder of the bearing lug to oppose the cutter seal groove,and having a circumferential, generally cylindrical wall and a generallyradial end wall; at least one rigid ring positioned between the shaftand cutter seal grooves with one sealing face engaging the opposedsealing face carried by a selected one of the shaft and the cutter, atleast one resilient energizer ring sealingly engaging a selectedcircumferential wall of the cutter and shaft seal groove and sealinglyengaging the rigid ring; the axial width of the rigid ring being lessthan the axial width of the shaft and cutter grooves when the cutter isthrust outward on the bearing shaft to define at least one axialclearance to permit unrestricted axial movement of the seal assemblybetween the radial end walls; the diameter of the journal bearing beingintermediate the circumferential walls of the cutter and shaft sealgrooves to decrease axial movement of the rigid ring with respect toaxial cutter movement to a selected ratio during drilling.
 8. Theinvention defined by claim 7 wherein the ratio of the seal movement toaxial cutter movement is selected to decrease seal movement, consistentwith the following formula:

    S=2G-C/H+C+2A

Where: S=Seal Movement For Cutter Movement Relative to the ShaftH=Effective Shaft O-Ring Annular Area C=Effective Cutter O-Ring AnnularArea A=Effective Rigid Ring Annular Area G=Effective Cutter Seal GrooveAnnular Area.
 9. The invention defined by claim 8 wherein each of theresilient energizer rings is of the O-ring type and the ratio of rigidring movement to axial cutter movement is substantially one half to one.10. An earth boring bit with an improved pressure compensating, rigidface seal assembly having one rigid ring and sealing face carried by abearing shaft and an opposed sealing face carried by the cutter, saidbit comprising:a body which includes at least one leg and a cantileveredand threaded bearing lug that extends downwardly and inwardly; alubrication system formed at least partially in the body, including ahydrostatic pressure compensator, to lubricate said bearings; thebearing lug including a threaded inner end and an outer base region witha transverse shoulder secured to the body; a bearing sleeve with asubstantially cylindrical journal bearing surface containing an assemblygroove for a cutter retainer means, a threaded interior secured to thethe threaded end of the bearing lug, and a mouth engaging the transverseshoulder of the lug; a rotatable cutter having an interior cylindricalbearing surface, a retainer groove that registers with the assemblygroove in the bearing sleeve, and an annular, cutter seal groove in theouter end of the bearing surface; retainer means between the assemblygroove and the retainer groove; a cutter seal groove formed near theoutermost region of the cylindrical bearing in the cutter to have acircumferential, generally cylindrical wall and a generally radial endwall; a shaft seal groove formed at least partially in the outer end ofthe bearing sleeve and the transverse shoulder of the bearing lug tooppose the cutter seal groove, and having a circumferential, generallycylindrical wall and a generally radial end wall; a pair of rigid ringspositioned in the seal grooves to have opposed, sealing faces; a pair ofresilient energizer rings, each of which sealingly engages a respectiveone of the rigid rings and one of the oppositely facing circumferentialwalls of the cutter and shaft seal grooves to define a seal assembly;the axial width of the rigid rings being less than the axial width ofthe shaft and cutter seal grooves when the cutter is thrust outward onthe bearing shaft to define at least one axial clearance to permitunrestricted axial movement of the seal assembly between the radial endwalls; the diameter of the journal bearing being intermediate thecircumferential walls of the cutter and shaft seal grooves to decreaseaxial movement of the seal assembly to a selected ratio when the cuttermoves axially on the bearing shaft during drilling.
 11. The inventiondefined in claim 10 wherein the ratio of the seal movement to axialcutter movement is selected to decrease seal movement, consistent withthe following formula:

    S=2G-C/H+C+2A

Where: S=Seal Movement For Unit Cutter Movement relative to the ShaftH=Effective Shaft O-Ring Annular Area C=Effective Cutter O-Ring AnnularArea A=Effective Rigid Ring Annular Area G=Effective Cutter Seal GrooveAnnular Area.
 12. The invention defined by claim 11 wherein each of theresilient energizer rings is of the O-ring type and the ratio of rigidring movement to axial cutter movement is substantially one half to one.13. An earth boring bit with an improved seal means and pressurecompensating system, said bit comprising:a body which includes at leastone leg and a cantilevered and threaded bearing lug that extendsdownwardly and inwardly; a lubrication system formed at least partiallyin the body, including a hydrostatic pressure compensator, to lubricatesaid bearings; the bearing lug including a threaded inner end and anouter base region with a transverse shoulder secured to the body; abearing sleeve with a substantially cylindrical journal bearing surfacecontaining an assembly groove for a cutter retainer means, a threadedinterior secured to the threaded end of the bearing lug, and a mouthengaging the transverse shoulder of the lug; a rotatable cutter havingan interior cylindrical bearing surface, a retainer groove thatregisters with the assembly groove in the bearing sleeve, and anannular, cutter seal groove in the outer end of the bearing surface;retainer means between the assembly groove and the retainer groove; acutter seal groove formed near the outermost region of the cylindricalbearing in the cutter to have a circumferential, generally cylindricalwall and a generally radial end wall; a shaft seal groove formed atleast partially in the end of the bearing sleeve and the transverseshoulder of the bearing lug to oppose the cutter seal groove, and havinga circumferential, generally cylindrical wall and a generally radial endwall; a pair of rigid rings positioned in the seal grooves to haveopposed and engaging sealing faces; a pair of resilient energizer rings,each of which sealingly engages a respective one of the rigid rings andone of the oppositely facing circumferential walls of the cutter andshaft seal grooves to define a seal assembly positioned between the endwalls of the cutter and shaft seal grooves; the seal assembly beingexposed to and moved axially by dynamic pressure differentials betweenthe lubricant and the ambient drilling fluids; the axial width of theengaged rigid rings and seal assembly being less than the axial, minimumwidth of the cutter and shaft seal grooves when the cutter is thrustoutwardly on the bearing shaft to define at least one axial clearance topermit axial movement of the rigid rings between the end walls of thegroove when the cutter moves relative to the bearing shaft; the diameterof the journal bearing being intermediate the circumferential walls ofthe seal groove to decrease axial movement of the seal assembly to aselected ratio when the cutter moves axially on the bearing shaft. 14.The invention defined by claim 13 wherein the ratio of the seal movementto axial cutter movement is selected to decrease seal movement,consistent with the following formula:

    S=2G-C/H+C+2A

Where: S=Seal Movement For Unit Cutter Movement Relative to the ShaftH=Effective Shaft O-Ring Annular Area C=Effective Cutter O-Ring AnnularArea A=Effective Rigid Cutter Annular Area G=Effective Cutter SealGroove Annular Area.
 15. The invention defined by claim 13 wherein eachof the resilient energizer rings is of the O-ring type and the ratio ofrigid ring movement to axial cutter movement is substantially one halfto one.